U.S. patent number 5,459,383 [Application Number 08/259,361] was granted by the patent office on 1995-10-17 for robust active damping control system.
This patent grant is currently assigned to Quantum Corporation. Invention is credited to Harry S. Hvostov, Michael D. Sidman.
United States Patent |
5,459,383 |
Sidman , et al. |
October 17, 1995 |
Robust active damping control system
Abstract
Methods and apparatus for providing improved gain and bandwidth
and robust damping of mechanical resonances for a servo control
system by providing a collocated or nearly collocated feedback loop
that operates in parallel with the ordinary feedback loop used for
position or velocity feedback.
Inventors: |
Sidman; Michael D. (Colorado
Springs, CO), Hvostov; Harry S. (Acton, MA) |
Assignee: |
Quantum Corporation (Milpitas,
CA)
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Family
ID: |
24617251 |
Appl.
No.: |
08/259,361 |
Filed: |
June 14, 1994 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
Issue Date |
|
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652554 |
Feb 7, 1991 |
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Current U.S.
Class: |
318/611; 318/560;
318/561; 318/618; 318/628; G9B/21.013 |
Current CPC
Class: |
G05B
19/19 (20130101); G11B 21/081 (20130101); G01B
21/045 (20130101); G05B 2219/37388 (20130101); G05B
2219/41025 (20130101); G05B 2219/41117 (20130101); G05B
2219/42077 (20130101) |
Current International
Class: |
G05B
19/19 (20060101); G11B 21/08 (20060101); G05B
005/01 () |
Field of
Search: |
;318/560-646
;360/70-88 |
References Cited
[Referenced By]
U.S. Patent Documents
Other References
Sidman, "Adaptive Control of a Flexible Structure," A Dissertation
Submitted to the Department of Electrical Engineering and the
Committee on Graduate Studies of Stanford University, Jun.
1986..
|
Primary Examiner: Ip; Paul
Attorney, Agent or Firm: Harrison; David B.
Parent Case Text
This application is a continuation of application Ser. No.
07/652,554 filed Feb. 7, 1991, now abandoned.
Claims
What is claimed is:
1. A method of stabilizing a control servo system including an
actuator system having a controlled element, an actuator for moving
said controlled element, and a mechanical structure coupling said
actuator and said controlled element, said servo control system
further including a non-collocated position sensor for sensing a
displacement of said controlled element, a controller for
positioning said controlled element by controlling said actuator, a
primary feedback loop from said position sensor to said controller,
comprising steps of:
embedding a secondary feedback loop within said primary feedback
loop, the secondary feedback loop including a substantially
collocated transducer positioned on said actuator system in an area
in proximity to said actuator;
converting mechanical oscillations of said actuator system at the
area in proximity to said actuator into a nearly collocated motion
signal;
processing said motion signal, said step of processing said motion
signal including the steps of filtering said motion signal to
accentuate said motion signal at a resonant frequency of said
actuator system, amplifying said filtered motion signal and
modifying the amount of amplification of said filtered motion
signal in response to the level of said filtered motion signal
relative to a reference signal level; and
modifying the effect of said controller in said primary feedback
loop on said actuator in response to said processed motion signal
from said secondary feedback loop to actively dampen oscillations
of said actuator system at said resonant frequency.
2. Apparatus for stabilizing a control servo system including an
actuator system having a controlled element, an actuator for moving
said controlled element and a mechanical structure coupling said
actuator and said controlled element, said control servo system
further including a non-collocated position sensor for sensing a
displacement of said controlled element, a controller for
positioning said controlled element by controlling said actuator
and a primary feedback loop from said position sensor to said
controller, comprising:
a secondary feedback loop embedded within said primary feedback
loop, the secondary feedback loop comprising:
a substantially collocated transducer positioned on said actuator
system in an area in proximity to said actuator for converting
mechanical oscillations of said actuator system in the area in
proximity to said actuator to a motion signal;
means for processing said motion signal, said means for processing
comprising means for filtering said motion signal to accentuate
signal components at a resonant frequency of said actuator system,
means for amplifying said filtered motion signal and limiting means
for modifying the amount of amplification of said filtered motion
signal in response to a reference signal level; and
means for modifying the effect of said controller in said primary
feedback loop on said actuator in response to said processed motion
signal from said secondary feedback loop to actively dampen
oscillations of said actuator system at said resonant
frequency.
3. Apparatus for stabilizing a control servo system including an
actuator system having a controlled element, an actuator for moving
said controlled element and a mechanical structure coupled with
said controlled element and said actuator for coupling said
actuator and said controlled element, said control servo system
further including a position sensor for sensing a displacement of
said controlled element, a controller foro positioning said
controlled element by controlling said actuator and a feedback loop
from said position sensor to said actuator for modifying the effect
of said controller on said actuator, by actively dampening both
internally and externally created mechanical resonances in said
actuator system, comprising:
means for converting said mechanical resonances of said actuator
system to a motion signal, said means for converting coupled to
said mechanical structure in proximity to the coupling of said
actuator to said mechanical structure;
means for processing said motion signal, said means for processing
comprising means for filtering said motion signal to accentuate
signal components representative of said mechanical resonances,
means for amplifying said filtered motion signal and limiting means
for modifying the amount of amplification of said filtered motion
signal in response to a reference signal level;
means for modifying the effect of said controller on said actuator
in response to said processed motion signal; and
wherein said limiting means comprises a magnitude comparator that
compares the level of said filtered motion signal with said
reference signal level to produce a gain select signal
representative of said comparison, and a retriggerable timer that
responds to said gain select signal to hold the level of said gain
select signal for a period that exceeds half of the period of the
lowest frequency of said mechanical resonances to be actively
damped.
4. The apparatus as set forth in claim 3, wherein said actuator is
of the linear motion type and said converting means comprises means
for sensing linear acceleration.
5. The apparatus as set forth in claim 3, wherein said actuator is
of the rotary motion type and said converting means comprises means
for sensing angular acceleration.
6. The apparatus as set forth in claim 5, wherein said means for
sensing angular acceleration comprises a pair of linear
accelerometers arranged in relative opposition to accentuate
tangential acceleration components while substantially rejecting
translational acceleration components of the actuator.
7. The apparatus as set forth in claim 3, wherein said converting
means comprises means for sensing velocity.
8. The apparatus as set forth in claim 3, wherein said converting
means comprises means for sensing mechanical strain.
9. The apparatus as set forth in claim 3, wherein said converting
means comprises means for sensing flexure.
10. A robust control servo system including an actuator system that
actively dampens resonances, said actuator system having a
controlled element, an actuator for moving said controlled element
and a mechanical structure coupling said actuator and said
controlled element, comprising:
a non-collocated feedback loop that comprises a position sensor for
sensing a displacement of said controlled element and a controller
for positioning said controlled element by controlling said
actuator; and
a nearly collocated feedback loop embedded within said
non-collocated feedback loop comprising means for converting said
mechanical resonances of said actuator system to a motion signal,
said means for converting coupled to said mechanical structure in
proximity to the coupling of said actuator to said mechanical
structure, means for processing said motion signal to accentuate
said motion signal in the frequency range of said mechanical
resonances of said actuator system and means for modifying the
effect of said controller on said actuator in response to said
processed motion signal.
Description
FIELD OF THE INVENTION
The present invention relates to electromechanical servo control
systems, and more particularly to methods and apparatus for
improving the response characteristics of the servo control systems
with an auxiliary feedback loop to provide robust active damping of
mechanical resonances.
BACKGROUND OF THE INVENTION
Electromechanical servo control systems as applied to
electromechanical positioning and velocity control systems are
often adversely affected by mechanical actuator resonances. These
resonances generally cannot be damped effectively by conventional
servo control systems. Typical servo systems that have mechanical
actuator resonances include those used for transducer head
positioning in disc drives. These mechanical actuator resonances
limit control loop gain of the servo system, reduce bandwidth of
the servo system, or both. This causes the controlled element, such
as a transducer head, to have excessive settling time after
positioning, poor response to disturbances, poor tracking ability,
or any combination of these.
Control loop stability problems may also result from these lightly
damped structural resonances that are associated with the
mechanical actuator. Prior art systems have made use of gain
stabilizing filters such as electronic notch filters inserted in
the control loop path. These filters are inserted in the forward
path of the control loop to filter out the signal information
within the band reject frequency range of the notch and therefore
help minimize excitation of these actuator resonances by the servo
control system itself.
The technique utilizing notch filters described above allows the
servo control system to effectively ignore lightly damped
structural actuator resonances in that very little control effort
may be applied by the servo controller at frequencies where notch
filters attenuate the control signal. However, the use of notch
filters does nothing to reduce the sensitivity of the servo system
to other types of disturbances that excite the mechanical
resonances, such as those caused by servo amplifier saturation and
distortion, external forces on the carriage such as caused by seek
activity, air turbulence, stiction, and so forth. This is because
such disturbances are typically generated at points in the control
which does not lend themselves to correction when such gain
stabilizing filters exist in the control loop. Although the notch
filters, coupled in a forward path configuration, will serve to
reduce steady state frequency components of the control signal in
the bandwidth of the notched structural resonances, they do not
necessarily inhibit the excitation of these resonances by these
disturbances external to the servo loop.
The exciting effect of power amplifier non-linearity, when due to
saturation or cross-over distortion, is generally not eliminated
with the notch filtering process, and such forms of disturbance,
appearing directly at the input to the mechanical process to be
controlled, continues to provide undesirable excitation of the
lightly damped structural resonances. Likewise, seek induced
excitation and other mechanical disturbances that are due to
external excitation other than that of the actuator itself are not
reduced by the notch filtering process.
Gain stabilization in the form of low-pass filtering is also used
for control loop gain stabilization. In this case, the cutoff
frequency of a low-pass filter that is inserted in the control loop
is generally lower than the frequencies of any of the lightly
damped resonances of the actuator structure that are desired to be
attenuated in their effect on the control loop. In this way, the
signal components of the control signal are substantially prevented
from exciting the lightly damped resonances of the actuator
structure. This helps ensure system stability, but it also
increases phase shift at frequencies in the vicinity of the servo
loop's unity gain crossing, thereby reducing the bandwidth of the
servo system. This is true of all gain stabilizing filters,
including notch filters. This reduction in bandwidth in turn
reduces the ability of the servo system to correct low frequency
vibration and tracking performance such as run out and other
disturbances that are due to external excitation and
non-linearities in positioning operations.
Typical servo control systems sense motion at so called
"non-collocated" locations on the actuator or its payload, such as
at arm tips on disc drives. "Non-collocated" in this context means
that there is substantial flexibility between the means of sensing
vibration and the "point of control" for the structure. The "point
of control" is the location on the actuator structure where control
effort is applied.
The prior art approaches to non-collocated control, as applied to
servo systems for transducer head positioning, cannot robustly
dampen structural resonances, such as the lightly damped motor
actuator resonances, because adequate loop gain cannot be
consistently maintained at or near resonant frequencies, especially
if there is variability in the magnitude or damping of actuator
resonances. This is because the non-uniform phase characteristic of
the actuator practically limits the loop bandwidth attainable with
closed loop stability, a mandatory requirement. For such systems,
the loop phase uncertainty becomes too high to close a stable high
gain feedback loop at or near resonant frequencies as required for
robust active damping. "Robust" means the control system is
relatively insensitive to changes in actuator dynamics, in this
case, resonances.
SUMMARY OF THE INVENTION
The present invention overcomes the ordinary limitations of
electromechanical positioning and velocity control systems such as
used for transducer head positioning in disc drives by utilizing a
collocated or nearly collocated feedback loop that is embedded
within the servo position feedback system loop. The collocated or
nearly collocated feedback loop has a motion sensor, such as an
accelerometer, that is located in the servo system at or near the
point of control.
Vibration of the structure caused by excitation both internal and
external to the control loop is robustly actively damped by the
system. Mechanical resonances excited by control action of the
actuator motor are damped by first measuring the vibrations at or
near the actuator motor itself before these vibrations are
propagated through the rest of the mechanical structure of the
actuator, and feeding back these measured vibrations with a high
gain feedback loop into the servo system so as to robustly actively
dampen them.
The collocated or nearly collocated feedback loop has the effect of
making the servo system perform as if the mechanical structure of
the system has much higher mechanical damping than it actually
possesses. Sufficient feedback is provided to damp out the
substantial resonances that adversely affect the servo system
within the frequency range necessary to provide the desired rate of
response. In this way, the servo system achieves high overall servo
loop gain over the entire frequency range of interest without
sacrificing system stability. System stability is provided by the
active damping control loop instead of notch or low pass filtering
the servo loop signal itself. The active damping loop compensator
is designed to impose much less phase loss than a servo loop that
is stabilized with low-pass and/or notch filters.
By increasing the system stability with the active damping control
loop, the servo system also achieves greater tolerance to system
component variations because the need to maintain the accurate
frequency and phase relationships in the feedback loop for the
notch filters are eliminated. Thus, the robustness, or tolerance of
the system to external and internal variation, is greatly
enhanced.
Lastly, the actively damped servo system, unlike a gain stabilized
servo system, can act to reduce the effect of excitation stemming
from external forces and disturbances on servo system positioning
excitation.
DESCRIPTION OF THE DRAWINGS
FIG. 1 is a diagram for a discrete mechanical model of a disc drive
system that incorporates an active damping system according to the
present invention.
FIG. 2 is a detailed diagram of a servo control system that
incorporates the active damping system according to the present
invention.
FIG. 3 is a preferred implementation of a gain scheduling feature
to provide automatic gain control for the active damping system
according to the present invention.
FIG. 4 is a graphical representation of a typical frequency
response magnitude characteristic for a non-collocated positioning
actuator mechanics of the servo control positioning system shown in
FIG. 2.
FIG. 5 is a graphical representation of a typical frequency
response magnitude characteristic for the nearly collocated
actuator mechanics of the servo positioning system shown in FIG.
2.
FIG. 6 is a graphical representation of the phase response for the
nearly collocated mechanics corresponding to FIG. 5.
FIG. 7 is a graphical representation of the phase response of the
non-collocated actuator mechanics corresponding to FIG. 4.
FIG. 8 is a graphical representation of a suitable frequency
response magnitude characteristic for the active damping
compensator according to the present invention.
FIG. 9 is a graphical representation of a suitable open loop
response magnitude characteristic of the active damping feedback
loop according to the present invention, comprising the motion
sensor response cascaded with the active damping compensator
response and the nearly collocated actuator response.
FIG. 10 is the phase response of the active damping loop
corresponding to FIG. 9.
FIG. 11 is a graphical representation of the closed loop
sensitivity frequency response magnitude characteristic of the
active damping loop.
FIG. 12 is the resultant closed loop non-collocated frequency
response magnitude characteristic of the noncollocated actuator
incorporating active damping according to the present invention as
a function of frequency.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
Referring to the drawings, wherein reference characters designate
like or corresponding parts throughout the views, FIG. 1 is a block
diagram of a typical rotary actuator disc drive system 2 that
incorporates an active damping system 4 according to the present
invention. Although the disk drive rotary actuator example
indicates how the active damping system is incorporated into a
motion control system, motion control applications for the active
damping system are not limited to disk drives, nor are they limited
to rotary actuators. The disc drive system 2 comprises a plurality
of actuator arm/head assemblies 6, with each of the assemblies 6
coupled to its own corresponding one of a plurality of actuator
shaft segments 8, each of which constitute an actuator shaft
position. The shaft segments 8 are in turn coupled to each other
and to an actuator motor 10.
For purposes of illustration, eight of the assemblies 6 are shown
coupled to eight of the shaft mountings 8 in FIG. 1, although the
disc drive system 2 may comprise a greater or lesser number of
these components, depending on data storage requirements. The
mechanical yaw motion flexibility and structural damping of each
assembly 6 is represented by a plurality of springs 12 and dash
pots 14 in FIG. 1.
The model of each assembly 6 is comprised of one of the springs 12
in parallel with one of the dash pots 14. Thus, for purposes of
illustration only, eight of the springs 12 are shown in combination
with eight of the dash pots 14 in FIG. 1.
Of course, each of the assemblies 6 is physically continuous or
distributed mechanical structure that does not physically comprise
one of the springs 12 in combination with one of the dash pots 14.
However, each of the assemblies 6 can be represented, for control
system modeling purposes, to have a characteristic spring constant,
represented by one of the springs 12, damped by a characteristic
damping pot represented by a corresponding one of the dash pots
14.
Consequently, each of the assemblies 6 have a characteristic
resonance with a frequency that is determined by the effective mass
of the corresponding actuator arm/head assembly 6 and the effective
spring constant of the assembly 6 represented by the associated one
of the springs 12. This characteristic resonance is damped in
amplitude by the structural damping of the coupling represented by
the associated one of the dash pots 14.
Likewise, each of the shaft mountings 8 are coupled to each other,
and the mountings 8 are in turn coupled to the actuator motor 10.
For purposes of illustration only, each of the shaft segments 8 is
represented by one of a plurality of springs 16 in parallel with a
corresponding one of a plurality of dash pots 18 that represent
torsional motion of each shaft segment 8.
Thus, each of the shaft segments 8 has its own characteristic
resonance that depends on the effective mass of any of the
assemblies 6 and shaft segments 8, and the shaft segment spring
constant represented by a corresponding one of the springs 16. The
characteristic resonance is damped in amplitude by the structural
damping of the shaft segments 8 represented by the associated one
of the dash pots 18.
Finally, the coupling between the shaft mountings 8 and the
actuator motor 10 is represented for purposes of illustration only
by a spring 20 in parallel with a dash pot 22. The shaft/motor
coupling has its own characteristic resonance that depends upon the
effective mass of the assemblies 6, shaft mountings 8 and actuator
motor 10 coupled to it and the spring constant represented by the
spring 20. The characteristic resonance is damped in amplitude by
the structural damping of the shaft/motor coupling represented by
the dashpot 22.
Because of the different spring constants and different effective
masses at different mechanical locations throughout the disc drive
system 2, it is evident that a plurality of different mechanical
resonant frequencies can be excited due to seek activity or
external disturbances of the disc drive system 2. Thus, the
mechanical resonances that may be excited in the disc drive system
2 extend over a range of frequencies. The actual set of resonant
frequencies and modes are actually due to the very complex
interaction of all the above-mentioned dynamics.
These resonance frequencies may be controlled to some extent by
passive design means to dampen the structure, or means to stiffen
the structure to raise them to frequencies above those of interest,
or both. However, practical design considerations usually prevent
these resonances from being effectively or completely controlled
with such measures. For instance, to secure higher mechanical
resonance frequencies in the disc drive system 2, it is necessary
to provide greater stiffness in the arms, shaft and shaft/motor
couplings.
Although greater stiffness can be achieved to some extent by
optimum design and selection of materials, it generally involves
increasing the size of the corresponding physical structure, and
therefore the mass. The greater mass however tends to lower the
frequency of resonance. Greater mass also increases the response
time of the disc drive system 2 for control purposes.
The resonances can be passively damped to a limited extent by
adding mechanical damping or structural damping materials, but the
increased inertia of the resultant physical structure then
decreases the response time of the disc drive system 2 for control
purposes and the variability of the resulting damping obtained over
temperature may be intolerable. Thus, the passive mechanical design
approaches for damping mechanical resonances in the disc drive
system 2 are often unsuited or only a partial solution.
The present invention reliably and effectively damps these
troublesome mechanical resonances in the feedback loop provided by
the active damping system 4. The active damping system 4 makes the
disc drive system 2 perform as if it has a much more rigid physical
structure for control purposes without increasing size or mass of
the physical structure and without sensitivity to ambient
conditions such as temperature and humidity.
The active damping system comprises a motion sensor 24, such as an
accelerometer, an active damping compensator 26, a summing element
28 and includes power amplifier 30. The motion sensor 24 is mounted
on or very near the source, or locus, of control effort for the
system. The locus of control effort is the point in the mechanical
structure of the system upon which the actuator motor 10 applies
control force or torque to the structure.
In this case, the motion sensor 24 is mounted on or very near the
actuator motor 10 to provide a collocated or nearly collocated
motion signal that represents the vibratory motion of the structure
that reflects excitation of the unwanted resonances. FIG. 1 shows
the motion sensor 24 at a nearly collocated location since as a
practical matter it is not possible to have sensor 24 in an ideal
collocated position which is illustrated by dotted lines and motion
sensor 24'. The motion sensor 24 is any suitable motion sensor,
such as an accelerometer, a velocity sensor, a strain gauge or a
flexure gauge. For sensing angular motion, the motion sensor 24 may
comprise a pair of linear accelerometers that are mounted in
opposition about the locus of the control effort to sense
tangential acceleration components and with limited sensitivity to
translational accelerational components.
The motion signal produced at the output of the motion sensor 24 is
fed to the input of the active damping compensator 26 via a motion
signal line 32. The active damping compensator 26 modifies the
motion signal received from the motion sensor 24 with selective
signal bandpass and amplification processing to produce an active
damping feedback signal.
The active damping compensator 26 processes the motion signal to
provide an active damping feedback signal that selectively
accentuates the frequency content of the sensed signal to coincide
with the range of frequencies of the resonances to be damped. This
amplifies the output of the motion sensor 24 proportional to the
amplitude of the resonances of the mechanical structure. The signal
processing functions of the active damping compensator 26 are
described in detail below and the details of the circuitry to
perform the functions are well known.
The active damping feedback signal produced at the output of the
active damping compensator 26 is fed to the input of the summing
element 28 via a line 34. The summing element 28 combines a command
signal on a line 36 from an associated servo controller (not shown)
in the disc drive system 2 with the active damping feedback signal
on the line 34.
The summing element 28 produces an actuator control signal that is
proportional to the command signal on the line 36 and the active
damping feedback signal on the line 34. The summing element 28 in
this case is a well known element and therefore not described in
further detail.
The actuator control signal produced at the output of the summing
element 28 is fed to the input of a power amplifier 30 via a line
38. The power amplifier 30 produces an actuator drive signal that
has a power level suitable for driving the actuator motor 10. The
power amplifier 30 typically is a transconductance type that
supplies the actuator motor 10 with current proportional to
actuator control signal voltage. The power amplifier 30 produces
the actuator drive signal by proportionally amplifying the actuator
control signal to the suitable level. The power amplifier 30 is a
well known element and therefore not described in further
detail.
The actuator drive signal produced at the output of the power
amplifier 30 is fed to the input of the actuator motor 10 via a
drive signal line 40. This completes the entire feedback loop for
the active damping system 4. Alternatively, if the actuator control
signal on the line 38 has a suitable power level, the power
amplifier 30 can be deleted, and the actuator control signal on the
line 38 can be used to drive the actuator motor 10 directly.
FIG. 2 shows a detailed block diagram of an electronic servo
control system 42 that uses the active damping system 4. The servo
system 42 comprises the actuator motor 10, the motion sensor 24,
the active damping compensator 26, the summing network 28 and the
power amplifier 30 as described above in connection with FIG.
1.
For purposes of illustration only, in FIG. 2 the mechanical
coupling of the servo feedback loop and the active damping loop is
represented as a non-collocated coupling 44 and a nearly collocated
coupling 46. The non-collocated coupling 44 represents the normal
actuator feedback transfer function(s) that provide(s) position
feedback to an ordinary servo controller 48. The nearly collocated
coupling 46 represents the transfer function of the mechanical
coupling of the motor 10 to the motion sensor 24 as described above
in connection with FIG. 1.
The servo controller 48 is a well known element that is not itself
part of the present invention, and therefore is not described in
further detail. The servo controller provides the command signal on
the line 36 as described above in connection with FIG. 1.
Thus, the servo control command signal and the active damping
feedback signal are summed together and amplified by the power
amplifier 30 to drive the actuator motor 10. The feedback provided
by the active damping compensator 26 thus contributes to the servo
control command signal to provide an actuator drive signal on the
line 40 that both reduces the unwanted vibrations sensed by the
motion detector 24 through the nearly collocated mechanical
coupling 46 and corrects position errors that are encountered
during positioning operations.
In the case of a servo position control system 42 used in a disc
drive, the arrangement and placement of the nearly collocated
coupling 46 and the motion sensor 24 depend on the operation of the
actuator motor 10 that is used. If the actuator motor is of the
voice coil or linear type, the motion that is sensed for the active
damping system 4 is preferably the linear acceleration of the
actuator motor 10.
To sense that portion of unwanted vibrations induced by the
actuator motor 10 that cause linear acceleration, the nearly
collocated coupling 46 is mounted to the carriage of the actuator
motor 10. The motion sensor 24 is preferably a single accelerometer
to measure linear motion transmitted through the nearly collocated
coupling 46.
If the actuator motor 10 is of the rotary type, the motion that is
sensed for the active damping system 4 is preferably the angular
acceleration of the actuator motor 10. This rotary motion is
preferably sensed by using a pair of accelerometers for the motion
sensor 24 and in this case the nearly collocated mounting 46
secures them in opposition to each other around the axis of
rotation of the actuator motor 10. Of course, other means for
sensing the rotary motion can be used without departing from the
teachings of this invention.
In either case, acceleration feedback, either directly sensed or
derived, is preferred to help elevate the magnitude of the sensed
mechanical resonances above the low frequency rigid body dynamics
in the collocated transfer function. Generally, active damping is
not required at low frequencies where the actuator behaves like a
rigid body.
Once the nearly collocated motion signal is available on the line
32, the active damping compensator 26 processes it into the active
damping loop feedback signal on the line 34 that is feedback to the
power amplifier 30 via the summing element 28. The command signal
on the line 36 from the servo controller 48 is combined with the
active damping loop feedback signal on the line 34 in the summing
element 28.
The drive signal on the line 38 actually comprises the combination
of the command signal on the line 36 and the active damping loop
feedback signal on the line 34. Thus, the active damping provided
by the present invention may be an add on system to an existing
servo control positioning system. The added components necessary
for active damping comprise the motion sensor 24, the active
damping compensator 26 and the summing element 28. The other
elements in the servo control positioning system 42 are those
normally used for servo control without active damping.
The transfer function from the command signal on the line 36 to the
displacement of the positioned elements, such as the transducer
heads located at the arm tips 6 in FIG. 1, is shown in FIG. 12 and
does not exhibit the lightly damped mechanical resonances that
occur without the active damping system 4 at frequencies where the
active damping compensator 26 passes or accentuates corrective
feedback signals. Without the active damping system, this transfer
function would appear as in FIG. 4, and contain lightly dampened
resonances.
The suppression of resonances achieved with the active damping
system 4 is completely different than the insertion of notch or low
pass filters in the control loop, as adopted in many servo control
positioning systems. With the active damping system 4, the gain of
the feedback loop through the nearly collocated signal path is
accentuated over the frequency range of the lightly damped
structural resonances.
With notch filter feed forward or low pass filter techniques, the
gain of the feedback loop is attenuated in the frequency range of
the lightly damped structural resonances. This means that there is
no effective loop gain at the resonant frequencies, and thus the
control loop expends no effort to damp the resonances. The notch or
low pass filtering techniques simply prevent the control loop from
exacerbating the instability of the system caused by presence of
such resonances.
Furthermore, the active damping system 4 is not acutely sensitive
to component parameters that may cause shifts in the frequencies or
damping of the lightly damped structural resonances or notch filter
bandstop frequencies. The high gain of the active damping system
extends over a wide range of frequencies that the system is
designed to actively dampen. The active damping system 4 also
introduces less phase shift than notch or low pass filtering of the
position feedback signal because the active damping system is
designed to have negligible loop gain at low frequencies.
The active damping system 4 is also superior because actively
damped actuator resonances are never allowed to be excited by seek
motion or other disturbances with the active damping system 4
turned on. Furthermore, the active damping system 4 may be left
operational in all modes of servo operation including times when
the actuator motor 10 is used for a seeking operation. Thus, the
active damper also operates while the servo controller is
regulating actuator velocity as well.
With active damping, the servo controller 48 is confronted with a
physical structure driven by the actuator motor 10 that appears
more like a rigid body over a greater frequency range. In this
sense the active damping system may be designed to be "invisible"
to the servo controller 48. This also allows the active damping
system 4 to be "plugged into" an existing servo system.
After the active damping system 4 is incorporated into the servo
positioning system 42, the servo controller 48 may be optimized for
higher control of gain or bandwidth. This is another major benefit
of incorporating the active damping system 4 into the servo control
positioning system 42. It enables significantly improved tracking
accuracy in the presence of run out, friction, windage and servo
controller quantization noise. In disc drives, the active damping
system 4 helps enable higher track densities.
The active damping compensator 26 is preferably operable over a
large range of active damping loop feedback signal amplitudes. To
achieve this, it is desirable that the active damping compensator
26 include an automatic gain control feature to prevent saturation
of the power amplifier 30 due to exceptionally large signal
amplitudes in the active damping feedback loop.
A preferred implementation of a gain scheduling feature to provide
automatic gain control for the active damping system according to
the present invention is shown in FIG. 3. It is described with two
discrete gain ranges that are automatically switched, although
additional gain ranges or a continuously variable gain range may be
used.
A gain scheduling active damping system 50 comprises the same
components as described above for the active damping system 4 in
connection with FIGS. 1 and 2. However, the gain scheduling active
damping system 50 also comprises a magnitude comparator 52 that has
a first input coupled to the active damping feedback loop signal
from the active damping compensator 26 on the line 34.
The magnitude comparator 52 compares the amplitude or magnitude
(its absolute value) of the active damping feedback loop signal
with a predetermined damping feedback signal limit level that is
coupled to a second input of the magnitude comparator 52 on a limit
level line 54. The level of the damping feedback limit level on the
line 54 is selected to correspond to a level of the active damping
feedback loop signal on the line 34 above which a change in active
damping feedback loop gain is necessary to ensure that saturation
of the power amplifier 30 does not occur due to active dampening
control action.
The magnitude comparator 52 simply provides an output that is
responsive to a difference in levels between its two inputs, and
therefore may constitute no more than an ordinary differential
amplifier. Its output is coupled to the input of a retriggerable
timer circuit 56 over comparator line 58. The retriggerable timer
56 produces a timer signal on a timer signal line 62.
The retriggerable timer 56 is triggered whenever the output signal
from the magnitude comparator 52 on the line 58 indicates that the
amplitude level of the active damping feedback loop signal on the
line 34 exceeds that of the feedback limit level on the limit level
line 54. Since the magnitude comparator 52 only senses
instantaneous signal levels, the purpose of the retriggerable timer
56 is to provide a constant loop gain for a preselected period
after the timer 56 is triggered, even when the active damping loop
signal on the line 34 only exceeds the level limit temporarily.
This may occur, for instance, when line 34 displays an oscillatory
characteristic and its level only exceeds the limit level on signal
peaks, but repetitively over the period required for correction. In
this case, it is desirable for the gain of the active damping
feedback loop to be reduced for a period that is sufficiently long
to let the active damping feedback loop signal decay to a level not
exceeding that of the limit level.
The retriggerable timer 56 provides this operation by generating a
constant level that indicates a gain selection command for a
predetermined period after the magnitude comparator 52 detects that
the active damping feedback signal exceeds the limit level. The
predetermined period is preferably selected to have a duration that
exceeds half of the period corresponding to the lowest frequency
actively damped mode.
The retriggerable timer 56 may comprise any timer circuit that
provides a constant output for a preselected period in response to
a trigger signal of selected amplitude. In this case, the trigger
signal for the retriggerable timer 56 is the output signal from the
magnitude comparator 52 on the line 58. The retriggerable timer 56
resets after the preselected period expires.
The output of the retriggerable timer 56 is fed to a gain select
input of a gain selector circuit 60 via a line 62. The gain
selector circuit 60 is preferably an amplifier that provides one or
another of two amplification levels of gain in response to the
output of the retriggerable timer 56 on the line 62. Thus, the
output of the timer 56 serves as a gain select signal for the gain
selector circuit 60.
A signal input for the gain selector circuit 60 is coupled to the
output of the active damping compensator 26. The output of the gain
selector circuit 60 is fed to one of the inputs of the summing
element 28 on a line 64.
The principle of the active damping system according to the present
invention calls for the use of dedicated local feedback with
accentuated loop gain in the frequency region or regions of poorly
damped electromechanical actuator resonances in an electronic servo
position control system such as used for transducer head
positioning in a disc drive. The crux of this invention is to
create a feedback loop structure, distinct from the positioning
servo system, whose main purpose is to provide substantial
auxiliary control action in the frequency range of the significant
resonances of the actuator used in the servo positioning system by
shaping the frequency response of the active damper feedback loop
in an appropriate manner.
To simplify the overall design for the active damping system, the
frequency response of the active damping system may be tailored to
have little or no gain at relatively low frequencies. This is
possible because all of the frequencies of any significance in the
active damper feedback loop are in a relatively high frequency
range. This approach allows the active damper feedback loop to be
designed, adjusted and operated independently of the associated
servo controller 48 because the different frequency ranges of
operation prevent them from adversely interacting with each
other.
The first step in designing the active damping compensator 26 for
one of the active damping systems 4 or 50 described above is to
measure or model the transfer function of the non-collocated
actuator in the servo positioning system. In this case, the
transfer function is represented in terms of the displacement of
the controlled element, such as one of the assemblies 6, as driven
by the actuator drive signal from the input of the power amplifier
30 on the line 38 as a function of the frequency of the actuator
drive signal.
A typical transfer function of a non-collocated actuator in the
servo control positioning system 42 as a function of frequency is
represented by a line 66 in FIG. 4. In this example, significant
actuator resonances are concentrated in the frequency range
extending from about 2 KHz to 3 KHz. To achieve active damping of
these resonances, the active damping motion sensor 24 must be
situated on the actuator structure so as to provide a significant
amount of signal in the frequency range of these resonances.
A typical frequency response characteristic for the nearly
collocated actuator mechanics as measured by motion sensor 24 is
represented by a line 68 in FIG. 5. The motion sensor 24 should
ideally be situated so that it is roughly coincident with the
location of actuating control effort to achieve the nearly
collocated phase response described below.
It is advantageous to select the mounting location for the motion
sensor 24 to be as close as possible to the point where control
force or torque is applied to the actuator structure. If this is
done in an ideal sense, the resultant collocated actuator transfer
function is distinguished by a pole/zero alternation pattern in the
S-plane. This pattern is very desirable for ensuring good stability
margins for the active damping feedback loop, especially when large
loop gain is introduced at high frequencies, where uncertainty in
the actuator dynamics may be problematical.
Generally, the phase response of a truly or ideally collocated
system is limited to a 180 degree range over the entire frequency
range. The phase response of the nearly collocated actuator
transfer function using one of the active damping systems 4 or 50
for the servo positioning system 42 in the example described above
is represented by a line 70 in FIG. 6. The phase shift for the
nearly collocated actuator transfer function represented by the
line 70 is limited to about a 220 degree range.
In contrast, the phase response of the non-collocated actuator
transfer function using the servo positioning system 42 without
active damping in the example described above is represented by a
line 72 in FIG. 7. The phase shift for the non-collocated actuator
transfer function represented by the line 72 extends over
approximately 370 degrees.
The much reduced phase shift of the nearly collocated actuator
transfer function represented by the line 70 in FIG. 6, compared
with the noncollocated phase shift represented by the line 72 in
FIG. 7, permits feedback loop design of greater stability with
higher overall servo loop gain, even at high frequencies. Thus,
high frequency attenuation with the use of notch or low pass
filtering is not required.
Of course, although the collocated transfer functions described
above are specifically directed to sensing linear or angular
acceleration, other transfer functions that are derived from other
sensed parameters, such as velocity, displacement or strain, may be
utilized in a similar fashion.
One objective of the active damping system according to the present
invention is to prevent the actuator resonances in an electronic
servo control positioning system from being excited by external
inputs, including control forces or torques and mechanical or
electrical disturbances to the actuator. To avoid such excitation
from these inputs, the active damping system must generate an
active damping loop feedback signal that serves to match the total
control and disturbance excitation to the actuator at the selected
resonant frequencies to counteract their effect.
In FIG. 2, this active damping feedback loop signal is produced by
the active damping compensator 26 on the line 34. Therefore, the
active damping compensator 26 must have access to the actuating
signal, the motion signal on the line 32 provided by the nearly
collocated motion sensor 24 in FIG. 2. Once the active damping
feedback loop is closed, by summing the active damping loop
feedback signal on the line 34 with the servo loop command signal
on the line 36 in the summing network 28, substantial reduction of
the effect of these disturbances takes place to minimize actuator
resonance excitation.
To this end, it is desirable to characterize the complementary
sensitivity transfer function of the active damping loop to have a
magnitude of approximately one in the frequency range, or ranges,
of the actuator resonances. The complementary sensitivity transfer
function in this case, according to feedback theory, is the open
loop gain divided by the open loop gain plus one. Thus, the open
loop gain of the active damping compensator 26 must be reasonably
high within the frequency range of the unwanted actuator
resonances.
In the example described above, the unwanted resonances, as shown
in FIG. 4, are primarily in the 2 KHz to 3 KHz range. A suitable
transfer function for the active damping compensator 26 as a
function of frequency in this case is represented by a line 74 in
FIG. 8.
The open loop gain of the active damping compensator 26 represented
by the line 74 as a function of frequency is shown, with increasing
frequency, to exceed unity, or 0 db, at about 1.0 KHz. Increasing
in frequency from this 0 db level, it has two peak levels of
approximately 30 db at about 1.5 KHz and 2.1 KHz respectively, with
a droop between these peaks down to approximately 15 db at about
1.7 KHz. Beyond the upper peak gain level at about 2.1 KHz, the
gain decreases with increasing frequency to drop back down to 0 db
at about 3.7 KHz.
The open loop gain of the active damping compensator 26 represented
by the line 74 as a function of frequency is not less than
approximately 15 db throughout the range of frequencies from about
1.3 KHz to about 2.7 KHz. The open loop magnitude response of the
motion detector 24 cascaded with the active damping compensator 26
as a function of frequency, using the response of the sensor 24
lumped with the transfer function of the nearly collocated
mechanics 46 shown in FIG. 5, is indicated by a line 76 in FIG. 9.
The line 76 thus represents the overall open loop gain of the
active damping loop described above.
The open loop gain of the active damping loop is at least unity
within the frequency range extending from approximately 1.7 KHz to
approximately 3.7 KHz, with a peak of about 38 db at approximately
2.5 KHz. The gain is about 10 db at the 2 and 3 KHz points of the
active damping loop gain transfer function. Since the complementary
transfer function for the active damping loop represents its loop
gain divided by its loop gain plus one, the complementary
sensitivity transfer function for the active damping loop is
approximately one throughout the 2 KHz to 3 KHz frequency
range.
As a practical matter, the gain of the active damping loop as shown
in FIG. 9 is limited by the permissible amount of phase shift
within the bandpass range of the loop. In FIG. 10, a line 78
represents the corresponding phase shift of the active damping loop
as a function of frequency. Within the frequency range that the
active damping loop gain exceeds unity, the phase shift of the
active damping loop does not exceed approximately 160 degrees. This
variation amounts to low and high frequency phase margins of
approximately 40 degrees if a 200 degree variation from a reference
phase is permissible for stability.
The closed loop sensitivity magnitude transfer function of the
active damping loop is represented by a line 82 as a function of
frequency in FIG. 11. The sensitivity transfer function in this
case, according to feedback theory, is the reciprocal of the open
loop gain plus one. The sensitivity magnitude transfer function
represented by the line 82 thus indicates the amount of additional
resonance attenuation provided by the active damping loop for the
non-collocated transfer function, as represented by the line 66 in
FIG. 4. This transfer function is what servo controller 48 "sees"
without the active damping system.
The resultant-damped non-collocated transfer function of the
actuator as a function of frequency is represented by a line 80 in
FIG. 12. This transfer function is the result of cascading the
sensitivity transfer function represented by a line 82 in FIG. 11
with the original noncollocated transfer function represented by
the line 66 in FIG. 4. The transfers function in FIG. 12, is what
servo controller 48 "sees" with the active damping system working.
The unwanted resonances in this example are attenuated
approximately 30 db.
The effect of the active damper loop at lower frequencies
preferably is small. This is done intentionally to minimize
interaction between the nearly collocated active damper loop and
the non-collocated servo positioning loop, and to allow the servo
controller 48 to be designed independently of the active damping
loop. The active damper loop is also designed to have low gain at
frequencies higher than the frequency range of the unwanted
resonances to minimize its reaction to unmodeled high frequency
dynamics and wide band noise produced by the nearly collocated
motion sensor 24.
Because the active damping loop gain is generally high at least at
some frequencies, the magnitude of the signal in the active damping
loop must be limited to avoid saturation of the power amplifier 30.
The gain selector circuit 60 for the active damping system 50 shown
in FIG. 3 limits the magnitude of the active damping loop feedback
signal by selecting an appropriate loop gain value based on the
magnitude of the excitation of the active damping compensator 26,
as described above.
Thus, there has been described herein methods and apparatus for
providing improved gain and bandwidth for ordinary servo controlled
positioning systems by providing a collocated or nearly collocated
feedback loop that operates in parallel with the ordinary feedback
loop used for position feedback. It will be understood that various
changes in the details, methodology, parts, and configurations
described above to explain the nature of the present invention may
be made by those skilled in the art within the principle and scope
of the present invention as expressed in the appended claims.
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